1. Field of Invention
The present invention relates generally to the braking of an internal combustion engine, specifically to engine braking apparatus with two-level pressure control valves.
2. Prior Art
It is well known in the art to employ an internal combustion engine as brake means by, in effect, converting the engine temporarily into a compressor. It is also well known that such conversion may be carried out by cutting off the fuel and opening the exhaust valve(s) at or near the end of the compression stroke of the engine piston. By allowing compressed gas (typically, air) to be released, energy absorbed by the engine to compress the gas during the compression stroke is not returned to the engine piston during the subsequent expansion or “power” stroke, but dissipated through the exhaust and radiator systems of the engine. The net result is an effective braking of the engine.
An engine brake (or engine retarder) is desirable for an internal combustion engine, particularly for a compression ignition type engine, also known as a diesel engine. Such engine offers substantially no braking when it is rotated through the drive shaft by the inertia and mass of a forward moving vehicle. As vehicle design and technology have advanced, its hauling capacity has increased, while at the same time rolling and wind resistances have decreased. Accordingly, there is a heightened braking need for a diesel-powered vehicle. While the normal drum or disc type wheel brakes of the vehicle are capable of absorbing a large amount of energy over a short period of time, their repeated use, for example, when operating in hilly terrain, could cause brake overheating and failure. The use of an engine brake will substantially reduce the use of the wheel brakes, minimize their wear, and obviate the danger of accidents resulting from brake failure.
There are different types of engine brakes. Typically, an engine braking operation is achieved by adding an auxiliary engine valve event called an engine braking event to the normal engine valve event. Depending on how the engine valve event is produced, an engine brake can be defined as:                (a) Type I engine brake—the engine braking event is produced by importing motions from a neighboring cam, which generates the so called “Jake” brake;        (b) Type II engine brake—the engine braking event is produced by altering existing cam profile, which generates a lost motion type engine brake;        (c) Type III engine brake—the engine braking event is produced by using a dedicated valve lifter for engine braking, which generates a dedicated cam (rocker) brake;        (d) Type IV engine brake—the engine braking event is produced by modifying the existing engine valve lift, which normally generates a bleeder type engine brake;        (e) Type V engine brake—the engine braking event is produced by using a dedicated valve train for engine braking, which generates a dedicated valve (the fifth valve) engine brake.        
The engine brake can also be divided into two big categories, i.e., the compression release engine brake (CREB) and the bleeder type engine brake (BTEB).
Conventional compression release engine brakes open the exhaust valve(s) at or near the end of the compression stroke of the engine piston. They typically include hydraulic circuits for transmitting a mechanical input to the exhaust valve(s) to be opened. Such hydraulic circuits typically include a master piston that is reciprocated in a master piston bore by a mechanical input from the engine, for example, the pivoting motion of the injector rocker arm. Hydraulic fluid in the circuit transmits the master piston motion to a slave piston in the circuit, which in turn, reciprocates in a slave piston bore in response to the flow of hydraulic fluid in the circuit. The slave piston acts either directly or indirectly on the exhaust valve(s) to be opened during the engine braking operation.
An example of a prior art CREB is provided by the disclosure of Cummins, U.S. Pat. No. 3,220,392 (“the '392 patent”), which is hereby incorporated by reference. Engine braking systems based on the '392 patent have enjoyed great commercial success. However, the prior art engine braking system is a bolt-on accessory that fits above the overhead. In order to provide space for mounting the braking system, a spacer may be positioned between the cylinder head and the valve cover that is bolted to the spacer. This arrangement may add unnecessary height, weight, and costs to the engine. Many of the above-noted problems result from viewing the braking system as an accessory to the engine rather than as part of the engine itself.
As the market for compression release-type engine brakes (CREB) has developed and matured, there is a need for design systems that reduce the weight, size and cost of such retarding systems. In addition, the market for compression release engine brakes has moved from the after-market to original equipment manufacturers. Engine manufacturers have shown an increased willingness to make design modifications to their engines that would increase the performance and reliability and broaden the operating parameters of the compression release-type engine brake.
One possible solution to the above problems is to integrate components of the braking system with the rest of the engine components. The most popular choice is to integrate the engine braking components into the engine rocker arm. The so called integrated rocker brake (IRB) devices can be found in the following U.S. Pat. Nos. 3,367,312, 3,786,792, 3,809,033, 5,564,385, 6,152,104, 6,234,143, and 6,253,730. The drawbacks of the integrated rocker brakes are the complexity and high moment of inertia due to the added engine braking components in the rocker arm, which may cause no-follow of the valve train components and other side effects on the engine performance during positive power operation.
Another engine component with integrated engine braking components is the valve bridge. One or more braking pistons can be placed in the valve bridge to form a variable valve lifter. The variable valve lifter usually contains a hydraulic linkage with lost motion means. There may be a gap in the valve lifter, for example, between the cam and the cam follower. When fluid, normally, engine oil, is supplied to the lost motion system, the valve lifter is expanded to take up the gap in the valve lifter so that the full motion from the cam is transmitted to the engine valves through the hydraulic linkage. On the other hand, if the fluid in the lost motion system is released, than the valve train will be contracted due to the gap in the valve lifter and some of the motion from the cam will be lost.
U.S. Pat. No. 5,829,397 discloses a system with a hydraulic piston in the valve bridge for controlling the amount of lost motion between an engine valve and a valve actuation means. A high speed trigger valve is used to quickly dump or supply fluid to the lost motion system so that the right amount of lost motion is accurately controlled. With such a high speed trigger valve, the continual variation of the engine valve lift is achieved. The lost motion system is operable for both engine positive power and engine braking modes of operation. However, such a full variable valve actuation (VVA) system is complex, expensive and prone to reliability issues due to the high speed trigger valve.
U.S. Patent Pub. No. 20050211206 discloses another lost motion system integrated into the valve bridge. However, a special “external” spring is needed to make the system work. The spring is mounted between the engine and the rocker arm to bias the rocker arm against a hydraulic piston into the valve bridge, so that a gap is formed between the overhead cam and the cam follower when the lost motion system is turned off. The gap is much larger than the normal valve lash, which increases the tendency of no-follow or impact of the valve train components. The special “external” spring needs to meet two conflicting requirements. First, the spring needs to be strong enough to prevent any no-follow of the valve train components even at the highest engine speed when the lost motion system is turned off. Second, because the hydraulic piston is loaded by the same spring, the spring needs to be weak enough to let the oil pressure overcome the spring force and lift up the hydraulic piston as well as the rocker arm to eliminate the gap between the cam and the cam follower when the lost motion system is turned on. The refill of the engine oil to the lost motion system could be slow due to the high spring force on the hydraulic piston, which may cause the system not fully actuated at high engine speeds. A compromise needs to be made to get the right size of the spring. However, such compromise is not ideal or even impossible when the moment of inertia of the valve train is too large, especially with the pushrod type of engines.
Another disadvantage associated with the above bridge lost motion system is that the sealing member of the resetting device is biased down against the seat by a spring, which may cause two potential problems. First, the sealing member will be impacted during both the engine braking operation (which is desirable) and the normal engine operation (not desirable). Second, the sealing member biased down against the seat by a spring keeps the control fluid sealed in the hydraulic piston chamber, which increases the potential of false start of the engine brake during the normal engine operation if there is no-follow, valve floating, excess oil leakage or other abnormal conditions.
One more challenge with the above bridge lost motion system and other integrated engine braking systems is that they may need a rather complicated system to provide two levels of oil supply pressure. The first level or lower level of oil supply pressure is for the lubrication or the hydraulic lash adjuster during the regular or positive power operation. U.S. Pat. Nos. 2,380,051, 3,140,698, 4,677,723, 4,924,821 and 5,150,672 disclosed different ways of putting one or more hydraulic pistons in the valve bridge for valve lash adjustment. The second level or higher level of oil supply pressure is for the lost motion operation. U.S. Patent Pub. No. 20070175441 uses two oil passages to supply oil, which has been widely used in the automobile industry, and may cause more oil consumption and oil pressure drop.
The flow control valve for supplying oil to an engine braking system is normally a 3-way solenoid valve, such as the one disclosed by U.S. Pat. No. 4,251,051, which has done a decent job for the traditional bolt-on engine brakes. However, there are a few drawbacks on this valve. First, the size of the valve is still too big, especially for the integrated engine braking systems. Second, the screwed-on installation may not fit on many engines where the solenoid terminals need to be specially oriented. Third, the drain port is on the bottom of the valve, while the outlet or high pressure port on the coil side, which may cause oil leakage into the coil structure on top of the flow control valve. Also, the area on the ball exposed to high pressure is too large, which requires high spring force to retain the ball and high magnetic force to actuate the valve.
U.S. Pat. No. 5,477,824 discloses a flow control valve combining the function of a traditional 3-way solenoid valve and that of a one-way check valve, trying to reduce the size and complexity of the engine braking system. However, the valve has not found commercial application because a 6 cylinder engine would need 6 new solenoid valves while only one or two of the traditional solenoid valves are enough to meet the need of the 6 cylinder engine braking. Since the solenoid valve is the most expensive and the least reliable component on an engine braking system, more solenoid valves are not desirable. Another drawback of the above solenoid valve is that the high pressure acting on the bottom of the valve causes high up-lift force on the valve, which requires high hold-down or clamping force on the valve.
It is clear from the above description that the prior-art engine brake systems have one or more of the following drawbacks:                (a) The system has high moment of inertia that may affect normal engine performance;        (b) The system causes higher no-follow tendency;        (c) The system needs a special “external” spring;        (d) The system may not be fully actuated at high engine speeds due to the force of the special “external” spring on the hydraulic piston;        (e) The system needs extra space to mount the special “external” spring;        (f) The system does not have a 2-level pressure supply solenoid valve;        (g) The system needs a leakage free and low clamping force solenoid valve; and        (h) The system needs a more compact and orientation free solenoid valve.        